Reciprocating high-pressure pumps (commonly called frac pumps) are often used in oil and gas fields for hydraulic fracturing of rock formations to increase hydrocarbon yields. Such pumps are often truck-mounted for easy relocation from well-to-well. And they are usually designed in two sections: the (proximal) power section (herein “power end”) and the (distal) fluid section (herein “fluid end”). Each pump fluid end comprises at least one (and typically three or more) subassemblies, with each subassembly comprising a portion of (or substantially the entirety of) a pump fluid end subassembly housing (shortened herein to “pump housing”).
For each pump fluid end subassembly, its pump housing comprises a pumping chamber in fluid communication with a suction bore, a discharge bore, and a piston/plunger bore. A suction valve (i.e., a check valve) within the suction bore, together with a discharge valve (i.e., another check valve) within the discharge bore, control bulk fluid movement from suction bore to discharge bore via the pumping chamber. Pulsatile fluid flow through the pump results from periodic pressurization of the pumping chamber by a reciprocating plunger or piston within the piston/plunger bore. The resulting suction and pressure strokes alternately produce wide pressure swings in the pumping chamber (and across the suction and discharge valves) as the reciprocating plunger or piston is driven by the pump power end.
Such pumps are operated at peak pumped-fluid pressures in current practice up to about 15,000 psi, while simultaneously being weight-limited due to the carrying capacity of the trucks on which they are mounted. See, e.g., U.S. Pat. No. 7,513,759 B1, incorporated by reference.
Due to high peak pumped-fluid pressures, suction valves experience particularly wide pressure variations between a suction stroke, when the valve opens, and a pressure stroke, when the valve closes. For example, during a pressure stroke a valve body may be driven longitudinally toward contact with its corresponding valve seat with total valve closing force that may vary from about 50,000 to over 150,000 pounds (depending on pumped-fluid pressure and valve body transverse area). Valve-closure impact energy, in the form of a short-duration high-amplitude valve-closure impulse, is thus applied longitudinally to the proximal surfaces of the valve. Since little of this energy is dissipated in a conventional valve, it is necessarily transmitted (in the form of vibration) to the pump housing via the valve seat. As described below, the valve's vibration-generation effect is analogous to striking the valve seat repeatedly with an impulse hammer.
A valve-closure impulse is particularly prominent when it occurs as a conventionally-stiff valve body contacts a conventional frusto-conical valve seat. The valve body's longitudinal movement typically stops abruptly, together with the associated longitudinal movement of a proximal mass of pressurized fluid in contact with the valve body. The kinetic energy of the moving valve body and pressurized fluid is thus nearly instantly converted to a high-amplitude valve-closure impulse of short duration. The effect may be compared to that of a commercially-available impulse hammer configured to produce broad-spectrum high-frequency excitation (i.e., vibration) in an object struck by the hammer.
Thus, broad-spectrum high-frequency vibration predictably results from the high-energy valve-closure impulse characteristically experienced by a conventionally-stiff valve body contacting a conventional frusto-conical valve seat. This vibration is quickly transmitted via the valve seat, especially to proximate areas of the pump housing where it can be expected to excite damaging resonances that predispose the housing to fatigue failures. See, e.g., U.S. Pat. No. 5,979,242, incorporated by reference. Frac pump maintenance costs are well-known in the well service industry to be relatively high and growing, due to both rapid valve wear and the early emergence of structurally significant cracks suggestive of corrosion fatigue in the pump housing (particularly near the suction valve seat).
Proposed valve designs in the past have included relatively lighter valve bodies comprising lighter materials and/or one or more interior cavities. See, e.g., U.S. Pat. No. 7,222,837 B1, incorporated by reference. Notwithstanding the somewhat lower valve-closure impulse amplitudes theoretically associated with such lighter valve bodies, they have been less popular than heavier and substantially more rigid valve bodies. The latter valve bodies have historically been shown to be relatively durable, but that performance record was largely created in lower pressure applications where the vibration fatigue issues described above are less prominent.
The recent transition period from lower pressure pump applications to higher pressure applications generally might be compared to the transition from slow-turning two-cylinder automobile engines to higher-speed and higher-powered inline six-cylinder engines around the years 1903-1910. New engine failure modes became evident, though they were neither anticipated nor understood at the time. Whereas the earlier engines had been under-powered but relatively reliable, torsional crankshaft vibrations in the six-cylinder engines caused objectionable noise (“octaves of chatter from the quivering crankshaft”) and unexpected catastrophic failures (e.g., broken crankshafts). (Quotation cited on p. 13 of Royce and the Vibration Damper, Rolls-Royce Heritage Trust, 2003). The vibration problems, though never entirely eliminated, were finally reduced to manageable levels after several crankshaft redesigns and the development of crankshaft vibration dampers by Royce and Lanchester.
Analogously, new valve designs are needed now for reducing failures associated with valve-generated vibration. Repeatedly-applied valve-closure energy impulses must be modified to reduce the excitation of destructive vibration resonances in valves, pump housings, and related structures.